Spindle motor and disk drive furnished therewith

ABSTRACT

In a spindle motor utilizing dynamic-pressure bearings having a full-fill structure, a bearing configuration that balances and sustains at or above atmospheric pressure the internal pressure of the bearing oil. Thrust and radial bearing sections are configured within oil-filled bearing clearances in between the rotor, the shaft, and a shaft-encompassing hollow bearing member. A communicating passage one end of which opens on, radially inwardly along, the thrust bearing section is formed in the bearing member. Either axial ends of the bearing clearance in between the bearing member and shaft communicate through the passage. The communicating passage enables the oil to redistribute itself within the bearing clearances. Pressure difference between the axial upper and lower ends of the oil retained in between the bearing member and the shaft is compensated through the communicating passage, preventing incidents of negative pressure within the oil and of over-lift on the rotor.

BACKGROUND OF INVENTION

1. Technical Field

The present invention relates to spindle motors employing dynamic-pressure bearings in which oil is the working fluid, and to disk drives equipped with such spindle motors. The invention relates in particular to miniature, low-profile spindle motors that drive recording disks 2.5 inches and under, and to disk-drives equipped with such spindle motors.

2. Description of the Related Art

Dynamic-pressure bearings in which the fluid pressure of a lubricating fluid such as oil interposed in between the shaft and the sleeve is exploited in order to support the two letting the one rotate against the other have been proposed to date as bearings for spindle motors employed in disk drives that drive hard disks and like recording disks.

FIG. 1 depicts one example of a spindle motor employing dynamic-pressure bearings. This spindle motor in which conventional dynamic pressure bearings are employed is configured with a pair of axially separated radial bearing sections d, din between the circumferential surface of the motor shaft b, which is integral with the rotor a, and the inner peripheral surface of the motor sleeve c, into which the shaft b is rotatively inserted. Likewise, a pair of thrust bearing sections g, g is configured in between the upper surface of a disk-shaped thrust plate e that projects radially outward from the circumferential surface of the shaft b on one of its ends, and the flat surface of a step formed in the sleeve c, as well as in between the lower surface of the thrust plate e and a thrust bush f that closes off one of the openings in the sleeve c.

A series of micro-gaps is formed in between the shaft b and thrust plate e, and the sleeve c and thrust bush f, and oil as a lubricating fluid is retained continuously without interruption within these micro-gaps. The oil retained in the micro-gaps is exposed to the air only within a taper-seal area h provided at the upper-end opening (the other opening in the sleeve c) of the gap formed in between the circumferential surface of the shaft b and the inner peripheral surface of the sleeve c. (This sort of oil-retaining structure will be denoted a “full-fill structure” hereinafter.)

The dynamic-pressure bearings further include herringbone grooves d1, d1 and g1, g1 that are linked pairs of spiral striations formed in the radial bearing sections d, d and thrust bearing sections g, g. In response to the rotor a rotating the grooves d1, d1 and g1, g1 generate maximum dynamic pressure in the bearing-section central areas where the spiral striation links are located, thereby supporting loads that act on the rotor a.

With dynamic-pressure bearings in a full-fill structure, when the rotor a begins rotating, pumping by the dynamic-pressure-generating grooves dl, d1 and g1, g1 acts to draw oil in toward the center areas of each of the radial bearing sections d, d and thrust bearing sections g, g, peaking the fluid dynamic pressure in the bearing center areas; but the downside of this is that along the bearing edge areas the oil internal pressure drops. In particular, in response to the pumping by the dynamic-pressure-generating grooves d1, d1 and g1, g1 the internal pressure of the oil drops—falling below atmospheric pressure and eventually becoming negative—in the region between the pair of radial bearing sections d, d among where oil is retained between the circumferential surface of the shaft b and the inner peripheral surface of the sleeve c, and in the region adjacent the outer periphery of the thrust plate e located in between the thrust bearings g, g among where oil is retained surrounding the thrust plate e.

If negative pressure within the oil has been brought about, during such operations as when the bearings are being charged with oil for example, air will dissolve into the oil and appear in the form of bubbles. Sooner or later the bubbles will swell in volume due to elevations in temperature, causing the oil to exude outside the bearings. This leakage impairs the endurance and reliability of the spindle motor. Negative pressure occurring within the oil can also lead to the dynamic-pressure-generating grooves coming into contact with air bubbles, which invites vibration incidents and deterioration due to NRRO (non-repeatable run-out). Such consequences impair the rotational precision of the spindle motor.

If because some factor is off in the manufacturing process the radial clearance dimension of the micro-gap formed in between the inner peripheral surface of the sleeve and the circumferential surface of the shaft is formed wider at the lower end axially than at the upper end, then an imbalance in the pumping by the dynamic-pressure-generating grooves d1, d1 in the radial bearing sections d, d will arise, and in the oil retained in between the inner peripheral surface of the sleeve and the circumferential surface of the shaft the pressure on the upper end axially will become higher than the pressure on the lower end axially. Pressure is consequently transmitted from along the upper axial end to along the lower axial end of the micro-gap formed in between the inner peripheral surface of the sleeve and the circumferential surface of the shaft, raising higher than is necessary the internal pressure of the oil retained in between the thrust-plate undersurface and the thrust bush and producing over-lift on the rotor, lifting it more than the predetermined amount.

Incidents of over-lift on the rotor give rise to frictional wear on the thrust plate and the sleeve by bringing them into contact; frictional wear is one factor to blame for spoiling bearing endurance and reliability. What is more, in the case of spindle motors for driving hard disks, as a consequence of the scaling-up of hard disk capacity, hard-disk recording faces and magnetic heads are being arranged extremely close to each other, and thus over-lift on the rotor brings the hard disk and magnetic heads into contact, leading to disk crash.

SUMMARY OF INVENTION

An object of the present invention is to realize a miniature, low-profile spindle motor that nonetheless provides for stabilized rotation.

Another object of the invention is to realize a spindle motor capable of sustaining at or above atmospheric pressure the internal pressure of the oil retained within the bearing clearances, and preventing air bubbles from being generated within the oil.

Still another object is to realize a spindle motor that enables the internal pressure of the oil retained within the bearing clearances to balance.

A different object of the present invention is to realize a spindle motor in which while miniaturization and slimming in profile are practicable, air bubbles arising due to negative-pressure, and incidents of over-lift on the rotor can be prevented, and incidents of knocking and grazing in the bearing sections can be controlled.

The present invention is also the realization of a low-profile, low-cost disk drive providing for stabilized spinning of recording disks.

Yet another object of the invention is to realize a disk drive of superior reliability and endurance, capable of preventing incidents of read/write errors.

In one example of a spindle motor according to the invention, the rotor has a circular flat face extending radially outward from the circumferential surface of the shaft, and a series of bearing clearances filled with oil is formed in between the flat face of the rotor, and the shaft and a hollow cylindrical bearing member having a bearing hole into which the shaft is rotatively inserted. A thrust bearing section is formed in between the end face at an opening in the bearing member, and the flat face of the rotor; and a radial bearing section is formed in between the inner peripheral surface of the bearing hole and the circumferential surface of the shaft. In addition, a communicating passage that serves to balance the pressure within the bearing clearances is formed in the bearing member. One end of the communicating passage opens on the thrust bearing section radially inwardly therein, and either axial ends of a bearing clearance formed in between the inner peripheral surface of the bearing hole and the circumferential surface of the shaft communicate through the passage.

This configuration makes it possible to prevent incidents of negative pressure and over-lift in a spindle motor utilizing dynamic-pressure bearings having a full-fill structure.

The fact either axial end of the bearing clearance formed in between the inner peripheral surface of the bearing hole and the circumferential surface of the shaft communicate by means of the passage enables the oil to redistribute itself within the bearing clearances. Thus even in cases in which a pressure difference has occurred between the axial upper and lower ends of the oil retained in between the inner peripheral surface of the bearing member and the circumferential surface of the shaft, arising from dimensional margin of error in the dynamic-pressure-generating grooves provided in the radial bearing section, or because some factor is off in processing the inner peripheral surface of the bearing hole or the circumferential surface of the shaft, the pressure difference is compensated through the communicating passage, preventing incidents of negative pressure within the oil and of over-lift on the rotor.

Likewise, the fact that the compensation of the oil internal pressure takes place within a region that pressure-wise is sealed by the thrust bearing section cushions the drop in oil pressure that occurs when the motor is decelerating, which makes it possible to control knocking and grazing in the bearing sections, meaning that motor reliability and endurance are sustained to a high degree.

In one example of a disk drive according to the present invention is a recording-disk-spinning spindle motor in which the rotor has a circular flat face extending radially outward from the circumferential surface of the shaft, and a series of bearing clearances filled with oil is formed in between the flat face of the rotor, and the shaft and a hollow cylindrical bearing member having a bearing hole into which the shaft is rotatively inserted. A thrust bearing section is formed in between the end face at an opening in the bearing member, and the flat face of the rotor; and a radial bearing section is formed in between the inner peripheral surface of the bearing hole and the circumferential surface of the shaft. In addition, a communicating passage that serves to balance the pressure within the bearing clearances is formed in the bearing member. One end of the communicating passage opens on the thrust bearing section radially inwardly therein, and either axial end of the bearing clearance formed in between the inner peripheral surface of the bearing hole and the circumferential surface of the shaft communicate through the passage.

From the following detailed description in conjunction with the accompanying drawings, the foregoing and other objects, features, aspects and advantages of the present invention will become readily apparent to those skilled in the art.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a sectional view diagramming the configurational outline of a conventional spindle motor;

FIG. 2 is a sectional view diagramming the configurational outline of a spindle motor having to do with a first embodiment of the present invention;

FIG. 3A is an enlarged section view representing on a larger scale a portion of the sleeve configuration in the spindle motor illustrated in FIG. 2;

FIG. 3B is a fragmentary enlarged section view showing on a larger scale the configuration of the bearing sections in the spindle motor illustrated in FIG. 2;

FIG. 4 is a plan view representing the configuration of the thrust bearing section of the spindle motor illustrated in FIG. 2;

FIG. 5 is a sectional view diagramming the configurational outline of a spindle motor having to do with a second embodiment of the present invention;

FIG. 6 is an enlarged section view representing on a larger scale a portion of the sleeve configuration in the spindle motor illustrated in FIG. 5;

FIG. 7 is a plan view representing the configuration of the thrust bearing section of the spindle motor illustrated in FIG. 5; and

FIG. 8 is a sectional view schematically illustrating the internal configuration of a disk drive.

DETAILED DESCRIPTION

Embodiments according to the present invention of a spindle motor and of a disk drive equipped therewith will be explained in the following with reference to FIGS. 2 through 8, but the present invention is not limited to the embodiments set forth below. It will be appreciated that although for the sake of convenience in the description of the present embodiments, “upper/lower, above/below, etc.” are in the vertical direction of the drawings, the orientation of the spindle motor in its actually installed state is not limited.

I. First Embodiment

(1)Configuration of Spindle Motor

To begin with, a spindle motor in a first embodiment of the present invention will be explained with reference to FIGS. 2 through 4. The spindle motor in the first embodiment of the present invention includes: a rotor 2, composed of a rotor hub 2 a and a shaft 2 b provided coaxially with the rotational center of the rotor hub 2 a; a cylindrical housing 6 affixed into a circular boss part 4 a with which a bracket 4 is furnished; and a hollow similarly cylindrical sleeve 8 mounted within the housing 6. The circumferential margin of the rotor hub 2 a is furnished with a flange-shaped disk-carrying section 2 c on which recording disks (illustrated as disks 53 in FIG. 8) such as a hard disk are carried, and a yoke 10 is fitted along the undersurface of the disk-carrying section 2 c. A rotor magnet 12 is attached by adhesive or like means to the inner peripheral surface of the yoke 10. Also, a stator 14 radially opposing the rotor magnet 12 is affixed to the circumferential surface of the circular boss part 4 a.

The housing 6, shaped by press-working a sheet-metal material, is roughly in the form of a cup whose under side is closed over. Further, the sleeve 8 is formed from a porous, oil-containing sintered metal material in which copper powder, iron powder, or the like are sintered and impregnated with oil. Forming the sleeve 8 in this way from a porous, oil-containing sintered metal means that dynamic-pressure-generating grooves, which will be described later, may be formed in the sleeve 8 at the same time it is produced, reducing the manufacturing costs. Moreover, porous, oil-containing sintered metal, being superiorly lubricative, curtails the occurrence of abrasion to enhance bearing reliability and endurance.

A through-hole (bearing hole) is provided in the sleeve 8, axially piercing its core; the shaft 2 b is inserted in the through-hole. The circumferential surface of the shaft 2 b radially opposes the inner peripheral surface of the sleeve 8 across a gap, while the end face of the shaft 2 b axially opposes the inner surface of the closed-end portion 6 a of the housing 6 across a gap. The sleeve 8 is mounted so as to position the end face at its upper end at approximately the same height as the end face at the upper end of the housing 6, and so as to set the end face at its lower end in opposition to the inner surface of the closed-end portion 6 a of the housing 6 via a clearance. Likewise, the end faces at the upper ends of the housing 6 and the sleeve 8 axially oppose the circular face along the underside (flat face) of the rotor hub 2 a across a gap.

The gap formed in between the end faces at the upper ends of the housing 6 and the sleeve 8, and the face along the underside of the rotor hub 2 a; the gap formed in between the inner peripheral surface of the sleeve 8 and the circumferential surface of the shaft 2; the clearance formed in between the inner surface of the closed-end portion 6 a of the housing 6 and the end face of the shaft 2; and the clearance formed in between the end face at the lower end of the adjoining sleeve 8 and the inner surface of the closed-end portion 6 a of the housing 6 (each of these gaps/clearances, as well as clearances formed within communicating passages 9 that will be described shortly, taken together will be denoted “bearing clearances” hereinafter) are all consecutive. Oil is retained continuously without interruption within these consecutive clearances, wherein a full-fill structure is configured.

In addition, axial grooves 8 a are provided in the circumferential surface of the sleeve 8, extending from the end face at its upper end to the end face at its lower end. By attaching the thus-configured sleeve 8 to the inner peripheral surface of the housing 6 the communicating passages 9 are formed by means of the axial grooves 8 a and the inner peripheral surface of the housing 6. Oil is retained within the communicating passages 9 also, wherein the oil retained in the gap formed in between the inner peripheral surface of the sleeve 8 and the circumferential surface of the shaft 2 b runs successively through the gap formed in between the end faces at the upper ends of the housing 6 and the sleeve 8, and the face along the underside of the rotor hub 2 a, the clearance formed in between the end face at the lower end of the sleeve 8 and the inner surface of the closed-end portion 6 a of the housing 6, and the communicating passages 9. (How the oil is compensated through the communicating passages 9 will be described in detail later.)

The upper-end portion of the outer peripheral surface of the housing 6 is made into an annular flange 6 b that juts radially outward and is formed to have a sloping-face contour so that its outer peripheral surface shrinks diametrically with further separation from the upper-end face of the housing 6. Likewise, a portion of the rotor hub 2 a at the radially outer edge of the face along the underside is made into a peripheral wall 2 d that depends toward the bracket 4. The inner peripheral surface of the peripheral wall 2 d and the outer peripheral surface of the flange 6 b radially opposing and out of contact with each other.

By the outer peripheral surface of the flange 6 b being formed to have a sloping-face contour as just mentioned, the radial gap dimension of the gap defined in between the inner peripheral surface of the peripheral wall 2 d and the outer peripheral surface of the flange 6 b flares gradually toward the bracket 4 (in the direction of the distal edge of the peripheral wall 2 d); in other words, heading oppositely the gap has a tapered form. This means that a taper seal area 16 is configured by a functional association between the inner peripheral surface of the peripheral wall 2 d and the outer peripheral surface of the flange 6 b. Only in this taper seal area 16 does the oil retained in the above-described clearances meet the air, in an interface where the surface tension of the oil and atmospheric pressure balance, forming the oil-air interface into a meniscus.

The taper seal area 16 serves as an oil reserve whereby according to the amount of oil retained within the taper seal area 16, the position where the interface forms can shift to suit. This means that oil retained within the taper seal area 16 is supplied to the later-described bearing sections in response to a decrease in oil retention volume, and that oil that has increased volumetrically owing to thermal expansion or other causes is accommodated within the taper seal area 16.

The fact that a taper-shaped gap is thus formed in between the outer peripheral surface of the flange 6 b portion of the housing 6 and the inner peripheral surface of the peripheral wall 2 d portion of the rotor hub 2 a, constituting the taper-seal area 16 in which surface tension is exploited, means a diametrically larger taper-seal area 16, and a taper-seal area 16 whose axial dimension is relatively large. The capacity within the taper-seal area 16 is consequently enlarged to be able adequately to match even the thermal expansion of the large amount oil retained in dynamic pressure bearings of the full-fill structure.

An annular retaining ring 18 is affixed by adhesive or like means to the distal edge portion of the peripheral wall 2 d beyond the taper-seal area 16. The retaining ring 18 is fit snugly yet out of contact with the bottom of the flange 6 b, thereby constituting a structure that with respect to the housing 6 retains the rotor 2 from coming out.

The fact that the structure for retaining the rotor 2 is configured alongside the circumferential surface of the housing 6, means that the pair of radial bearing sections, which will be described in detail later, and the retaining structure, are not disposed ranged on the same line in the axial direction. This enables putting to use as a bearing the entirety of the axial height dimension of the mutually opposing outer circumferential surface of the shaft 2 b and inner peripheral surface of the sleeve, to realize further slimming down of the motor while maintaining bearing stiffness.

The top face of the retaining ring 18 and the bottom face of the flange 6 b, and the inner peripheral surface of the retaining ring 18 and the circumferential surface of the housing 6, are continuous with the taper-seal area 16, and oppose each other across a gap having a clearance dimension that is smaller than the minimum radial clearance dimension of the taper-seal area 16 gap.

Setting to be as small as possible the gap dimensions of the axial gap defined in between the top face of the retaining ring 18 and the bottom face of the flange 6 b, and the radial clearance formed in between the inner peripheral surface of the retaining ring 18 and the circumferential surface of the housing 6, enlarges the difference when the spindle motor is rotating between the flow speed of the air in these gaps in between the retaining ring 18 and the housing 6, and the flow speed of the air in the radial gap defined in the taper-seal area 16. The gap-defining surfaces of the housing 6, flange 6 b, and retaining ring 18 function as a labyrinth seal in which the difference in airflow speed makes greater the resistance against outflow to the bearing exterior of vapor that has arisen due to the oil gasifying, to keep vapor pressure of the oil in the vicinity of the boundary surface high, and that further prevents the oil from transpiring.

Thus providing the labyrinth seal in continuity with the taper-seal area 16 enables not only impeding the oil from flowing out as a fluid, but also blocking the outflow to the motor exterior of oil mist generated by the oil gasifying due to such causes as elevation in temperature of the motor external environment. This consequently prevents decline in the volume of oil retained, which can maintain stabilized bearing performance over the long term, and makes for bearings of high endurance and reliability.

(2) Configuration of Bearing Sections

Reference is made now to FIG. 3A, a sectional view of the sleeve 8. As illustrated in FIG. 3A, herringbone grooves 20 a for inducing fluid dynamic pressure in the oil when the rotor 2 spins, constituted by linking pairs of spiral striations that slope into each other from mutually opposing directions with respect to the rotary direction, are formed in the inner peripheral surface of the sleeve 8 along its upper end. An upper radial bearing section 20 is thus configured in between the inner peripheral surface of the sleeve 8 where the herringbone grooves 20 a are formed, and the circumferential surface of the shaft 2 b.

In the herringbone grooves 20 a of the upper radial bearing section 20, the spiral striation set located along the upper side is formed greater in axial dimension than is the spiral striation set located along the lower side. The herringbone grooves 20 thus are formed to generate, in response to the rotor 2 rotating, a dynamic pressure maximum in a locus biased downward from the center, and at the same time to produce pressure that presses in on the oil downward. This inward-pressing pressure keeps the internal pressure of the oil retained within the gap where located lower than the upper radial bearing section 20 at atmospheric pressure or more.

Likewise, herringbone grooves 22 a for inducing fluid dynamic pressure in the oil when the rotor 2 spins, configured by linking pairs of spiral striations that slope into each other from mutually opposing directions with respect to the rotary direction, are formed in the inner peripheral surface of the sleeve 8 along its lower end, wherein in between that inner peripheral surface and the circumferential surface of the shaft 2 b a lower radial bearing section 22 is configured.

The herringbone grooves 22 a formed in the lower radial bearing section 22 are designed so that the spiral striations generate a substantially equivalent pumping force—so that the groove fundamentals, which are axial dimension and inclination angle with respect to the rotary direction, or groove width and depth, will be the same. In other words, the spiral striations are configured to be linearly symmetrical with respect to where they link. In the lower radial bearing section 22 the greatest dynamic pressure consequently appears in the axially central part of the bearing.

Reference is now made to FIG. 4, which depicts pump-in spiral grooves 24 a that induce pressure heading radially inward (toward the shaft 2 b) in the oil when the rotor 2 spins, formed in the end face of the housing 6 at its upper end, wherein a thrust bearing section 24 is configured in between that end face and the face along the underside of the rotor hub 2 a.

It should be understood that the end face along the free-end portion of the shaft 2 b and the inner surface of the closed-end portion 6 a of the housing 6 function as a hydrostatic bearing section exploiting oil internal pressure heightened, as will later be described in detail, by the spiral grooves 24 a of the thrust bearing section 24.

(3) Shaft Support Method

The way in which the bearing sections configured as set out above function for journal support will next be described in detail.

Pumping force by agency of the herringbone grooves 20 a and 22 a in the upper and lower radial bearing sections 20, 22 rises attendant on rotation of the rotor 2, producing fluid dynamic pressure. As far as pressure distribution in the upper and lower radial fluid dynamic bearings 20, 22 is concerned, pressure rises abruptly from alongside either ends of the herringbone grooves 20 a, 22 a and becomes maximal where the spiral striations link. The fluid dynamic pressure generated in the upper and lower radial hydrodynamic bearings 20, 22 is utilized to support the rotor 2 through the upper/lower axial ends of the sleeve 8 and the shaft 2 b, and plays roles both in centering the rotor 2 and restoring it from deviations.

Radially inward-heading pressure is induced in the oil in the thrust bearing section 24 by the pump-in spiral grooves 24 a attendant on rotation of the rotor 2. The radially inward-heading pressure raises the oil internal pressure of the oil, generating fluid dynamic pressure acting in a lifting direction on the rotor 2, and meanwhile keeps the pressure of the oil that as a whole is retained deeper into the bearing clearances (more toward the closed end 6 a) than the thrust bearing section 24 at positive pressure. It should be understood that the fluid dynamic pressure induced in the thrust bearing section 24 does not rise abruptly as is the case with the upper and lower radial hydrodynamic bearings 20, 22; rather, at maximum it is at a level exceeding atmospheric pressure to a certain degree. The oil retained in the bearing clearances deeper inward than the thrust bearing section 24 is pressure-wise brought into an essentially sealed state by the radially inward-heading pressure generated in the thrust bearing section 24.

With the herringbone grooves 20 a formed in the upper radial bearing section 20 being axially asymmetrical in form, dynamic pressure that presses downward on the oil is generated, whereby dynamic pressure that becomes maximal in the locus biased a certain extent from the center of this bearing section toward the lower radial bearing section 22 is generated. This dynamic pressure supports the shaft 2 b across its axially upper side, and keeps the pressure in the region between the upper radial bearing section 20 and the lower radial bearing section 22 at positive pressure—atmospheric pressure or greater—to prevent negative pressure from arising.

Now the pressure generated in the thrust bearing section 24 is as just noted at a level somewhat in excess of atmospheric pressure, but for this alone to put sufficient lift on the rotor 2 would be problematic. Nevertheless, as described above the internal pressure of the oil in being retained in between the end face of the shaft 2 b and the inner surface of the housing 6 closed-end portion 6 a functions as a hydrostatic bearing section because the pressure is transmitted through the communicating passages 9 so as to become equal to the internal pressure of the oil raised by the fluid dynamic pressure induced in the thrust bearing section 24. These thrust-bearing 24 and hydrostatic bearing sections operate associatively to enable sufficient lift to be put on the rotor 2.

In the same regard, an annular thrust yoke 26 made of a ferromagnetic material is disposed in a position on the bracket 4 opposing the rotor magnet 12, generating magnetic attraction in the axial direction between the rotor magnet 12 and the thrust yoke 26. This magnetic force balances the lifting pressure on the rotor 2 that is generated by the thrust bearing section 24 and by the hydrostatic bearing section between the end face of the shaft 2 b and the inner surface of the housing 6 closed-end portion 6 a, stabilizing the thrust-oriented support of the rotor 2. This sort of magnetic urging on the rotor 2 can be effectuated also by, for example, displacing the magnetic centers of the stator 14 and the rotor magnet 12 axially from each other.

(4) Configuration and Function of Communicating Passages

The axial grooves 8 a provided in the circumferential surface of the sleeve 8 can be formed so that in sectional outline they are semicircular as shown in FIG. 4, or else are roughly rectangular or triangular, by die-stamping the sleeve 8 at the same time it is formed of a porous, oil-containing sintered metal into its cylindrical shape. However, the axial grooves 8 a can be formed by machining the sleeve 8 after it has been formed into its cylindrical shape.

As illustrated in FIG. 3B, the communicating passages 9, which continue from the upper end to the lower end of the sleeve 8 in the axial direction, are defined between the inner peripheral surface of the housing 6 and the axial grooves 8 a when the sleeve 8 is attached to the inner peripheral surface of the housing 6. Within the communicating passages 9, oil is retained as described earlier in continuity with the oil retained within the series of bearing clearances. By the same token, the internal pressure of the oil retained within the communicating passages 9 is balanced with the internal pressure of the oil retained within the bearing sections.

By either the micro-gap formed in between the inner peripheral surface of the sleeve 8 and the circumferential surface of the shaft 2 b, where the upper and lower radial bearing section 20 and 22 are configured, remaining consistent in its predetermined dimension, or by the herringbone grooves 20 a and 22 a remaining consistent in their predetermined precision, the oil retained in the bearing sections will be on par at least with the pressure generated in the thrust bearing section 24, meaning that the oil internal pressure will not go negative.

On the other hand, if due to discrepancy in processing the inner peripheral surface of the sleeve 8 or the circumferential surface of the shaft 2 b the micro-gap formed in between the inner peripheral surface of the sleeve 8 and the circumferential surface of the shaft 2 b is formed wider at its upper end axially than at its lower end, the dynamic pressure generated along the lower radial bearing section 22 will exceed the dynamic pressure generated in the upper radial bearing section 20, producing a flow of oil heading from along the axial lower side to along the upper side, risking that the internal pressure of the oil retained along the closed-end portion 6 a of the housing 6, i.e., deeper within the bearing clearances, will turn negative. Likewise, should the micro-gap formed in between the inner peripheral surface of the sleeve 8 and the circumferential surface of the shaft 2 b be formed narrower at its upper end axially than at its lower end, the dynamic pressure that the herringbone grooves 22 a provided in the upper radial bearing section 20 generate will go over the predetermined pressure, giving rise to negative pressure between the end face of the shaft 2 b and the closed-end portion 6 a of the housing 6. And when in this case the oil is set flowing from along the axial upper side to along the lower side, there would be a concern lest the internal pressure of the oil in between end face of the shaft 2 b and the closed-end portion 6 a of the housing 6 rise higher than is necessary and produce over-lift on the rotor 2.

In contrast to these scenarios, by providing the communicating passages 9, though the dynamic pressure generated in the thrust bearing section 24 declines somewhat, the attenuation will be transmitted to the oil retained along the closed-end portion 6 a of the housing 6, and therefore under normal conditions the internal pressure of the oil in that region will not go negative.

As mentioned earlier, the herringbone grooves 20 a provided in the upper radial bearing section 20, inasmuch as they are axially asymmetrical in form, generate dynamic pressure that presses downward on the oil to keep the pressure in the region between the upper radial bearing section 20 and the lower radial bearing section 22 positive at atmospheric or greater, preventing the occurrence of negative pressure. Meanwhile, the herringbone grooves 20 a generate compressive force that constantly sets the oil flowing, and that pressurizes the oil so as to recirculate it toward the upper radial bearing section 20, setting up a series of circulatory paths in which the oil is sent: from the lower radial bearing section 22 and from between the end face along the lower end of the sleeve 8 and the inner surface of the closed-off end portion 6 a of the housing 6, by way of the communicating passages 9 and by way of the interval between the end face along the upper end of the sleeve 8 and the face along the underside of the rotor hub 2 a, toward the axial upper-end areas of the circumferential face of the shaft 2 b and the inner peripheral surface of the sleeve 8.

The series of circulatory paths serves to balance the pressure of the oil within the bearing clearances by constantly setting it in motion in a given direction. This prevents occurrences of air bubbles due to negative pressure and occurrences of over-lift on the rotor 2, and markedly broadens the margin of error tolerated in the manufacturing process, thereby improving yield rate. For that matter, even should a factor in the manufacturing process be so far off that the fluid dynamic pressure generated along the lower radial bearing section 22 exceeds at its greatest pressure the inward-pressing pressure generated in the upper radial bearing section 20, the pressure difference would set the oil flowing in the direction opposite to that noted above, likewise setting up circulatory paths whereby the pressure difference would be canceled.

It will be appreciated that inasmuch the communicating passages 9 are disposed so that on the one end the passages 9 open radially inward of the thrust bearing section 24, oil pressure within regions of higher than atmospheric pressure is held constant. This means that the bearing sections where they are further inward than the thrust bearing section 24 are pressure-wise brought into a sealed state by the thrust bearing section 24.

If for instance the communicating passages 9 on the one end were to open in between the bearing sections and the taper seal area, as long as the predetermined dynamic pressure were generated in the bearing sections, such as when motor rotation is steady, sufficient supporting stiffness would be produced, and the likelihood that knocking or grazing in the bearing sections would arise would therefore be slight. If, however, the motor rotational speed were to drop, such as when the motor is halted, inasmuch as the communicating passages 9 on the one end opened into an area apart from the region sealed pressure-wise—i.e. into a region where the oil pressure would be equal to or otherwise below atmospheric pressure—the oil pressure that within the bearing section had been sustained high would drop abruptly, on account of the pressure difference with the oil pressure in the area where the communicating passages 9 would open.

By the pressure within the bearing sections thus dropping abruptly, the rotor 2 would be prone to wobbling or running eccentrically, meaning that knocking and grazing between parts such as the shaft 2 b and the sleeve 8 that constitute the bearing sections would arise. While conceivable causes for this include weight imbalance in the rotor 2 incorporating the recording disks carried by the rotor hub 2 a, processing and assembly tolerances in the parts composing the motor, or imbalance in magnetic force acting in between the stator 14 and the rotor magnet 12, with such knocking and grazing in the bearing sections reoccurring every time the motor is halted, the striking wear and tear on the parts composing the bearing sections would degrade the motor reliability and durability.

In contrast to this scenario, by having the communicating passages 9 open radially inward of the thrust bearing section 24, the pumping by the spiral grooves 24 a that induces radially-inward-acting fluid dynamic pressure in the oil will continue operating until just before the motor comes to a complete halt. Since the thrust bearing section 24 thus functions as a partition wall pressure-wise, pressure drop within the bearing sections is eased, and knocking and grazing of the parts that make up the bearing sections is mitigated, which holds degradation in motor reliability and durability in check.

II. Second Embodiment

With reference to FIGS. 5 through 7, explanation concerning a spindle motor in a second embodiment of the present invention will be made.

(1)Configuration of Spindle Motor

In the foregoing first embodiment, a spindle motor was described having a configuration in which a porous sintered metal is utilized as the material for the sleeve 8, and in which the communicating passages 9 are formed by the combining of the sleeve 8 and the housing 6. The spindle motor that will be described in the following in the second embodiment has a configuration in which the sleeve is formed from an aluminum alloy, a copper alloy, or stainless steel, thereby rendering the housing set out in the first embodiment unnecessary. The fact that the spindle motor in the second embodiment thus does not require the housing makes it possible to curtail the number of motor parts, and to curtail the number of motor assembly steps.

The spindle motor illustrated in FIG. 5 is equipped with: a rotor 106 constituted from a rotor hub 102—which is made up of a an approximately disk-shaped upper wall portion 102 a (top plate) and a cylindrical peripheral wall portion 102 b (cylindrical wall) depending downward from the outer rim of the upper wall portion 102 a—and from a shaft 104 one end of which perimetrically is fixedly fitted into the central portion of the upper wall portion 102 a of the rotor hub 102; a hollow cylindrical sleeve 108 rotatively supporting the shaft 104; a cover member 110 that closes over the lower portion of the sleeve 108 and opposes the end face of the shaft 104 at its free end; and a bracket 114 formed unitarily with a cylindrical boss portion 112 into which the sleeve 108 is fitted. A stator 116 having a plurality of teeth is arranged on the circumferential surface of the boss 112. Then a rotor magnet 118 is affixed to the inner peripheral surface of the peripheral wall portion 102 b of the rotor hub 102, wherein the rotor magnet 118 from its radially inward side opposes the stator 116 across a gap. Likewise, a flange-shaped disk-carrying portion 102 c by which recording disks such as a hard disk (illustrated as disks 53 in FIG. 8) are carried is provided on a lower end portion of the circumferential surface of the peripheral wall portion 102 b of the rotor hub 102. A hole 104 a standing in axial extension is formed in the shaft 104 coaxially with its rotational center axis. Male threads (not illustrated) for fastening a clamp (not illustrated) in order to retain the recording disks on the disk-carrying portion 102 c of the rotor hub 102 are clinched in the opening of the through-hole 104 a on the rotor-hub 102 side. Explanation of the configuration of the taper-seal area and of the retainer for the rotor 106 in the spindle motor of the second embodiment will be omitted here because they are substantially the same as those of the first embodiment.

(2) Configuration of Bearing Sections

A series of micro-gaps is formed in between the upper-end face of the sleeve 108 and the undersurface of the upper wall portion 102 a of the rotor hub 102; and—continuing from the upper wall portion 102 a of the rotor hub 102—in between the circumferential surface of the shaft 104 and the inner peripheral surface of the sleeve 108; and continuous therewith, in between the shaft 104 lower end face, and the inner surfaces of the sleeve 108 and cover member 110, wherein oil is retained continuously without interruption throughout these micro-gaps.

Herringbone grooves 122 a as represented in FIG. 6, which induce fluid dynamic pressure in the oil when the rotor 106 spins, are formed on the inner peripheral surface of the sleeve 108 alongside its upper-end face (alongside the rotor-hub 102 end). The herringbone grooves 122 a are constituted by linking pairs of spiral striations inclining into each other from mutually opposing directions with respect to the rotary direction. An upper radial bearing section 122 is thus configured between the inner peripheral surface of the sleeve 108 where the grooves 122 a are formed, and the outer circumferential surface of the shaft 104. Likewise, herringbone grooves 124 a that induce fluid dynamic pressure in the oil when the rotor 106 spins are formed on the inner peripheral surface of the sleeve 108 alongside the bottom end of the shaft 104 (alongside the cover member 110). The herringbone grooves 124 a are constituted by linking pairs of spiral striations inclining into each other from mutually opposing directions with respect to the rotary direction. A lower radial bearing section 124 is thus configured between the inner peripheral surface of the sleeve 108 where the grooves 124 a are formed and the outer circumferential surface of the shaft 104.

In this configuration, of the spiral striations that constitute the herringbone grooves 122 a formed in the upper radial bearing section 122, those spiral striations located alongside the upper end of the sleeve 108 are designed to be longer in axial dimension than the spiral striations located alongside the lower radial bearing section 124. Likewise, of the spiral striations that constitute the herringbone grooves 124 a formed in the lower radial bearing section 124, those spiral striations located alongside the lower end of the sleeve 108 are designed to be longer in axial dimension than the spiral striations located alongside the upper radial bearing section 122. This configuration makes it so that maximum pressure in the upper and lower radial bearing sections 122 and 124 appears in loci that are biased toward the midsections of the upper and lower radial bearing sections 122 and 124 where the spiral striations links are located, rather than toward the longitudinal centers of the bearing sections.

In addition, pump-in spiral grooves 126 a as represented in FIG. 7, which induce radially inward-heading (toward the shaft 104) pressure in the oil when the rotor 106 spins, are formed on the sleeve 108 upper-end face (face axially opposing the upper wall portion 102 a). A thrust bearing section 126 is thus configured between the upper-end face of the sleeve 108 and the undersurface of the upper wall portion 102 a of the rotor-hub 102. Furthermore, a hydrostatic bearing 128 exploiting oil internal pressure heightened, in a like manner as in the spindle motor of the foregoing first embodiment, by the spiral grooves 126 a of the thrust bearing section 126 is constituted in between the inner surface of the cover member 110 and, axially opposing it, the bottom-end face of the shaft 104. The way in which the bearing configuration of the spindle motor in the second embodiment functions for journal support is also in a like manner as in the spindle motor of the foregoing first embodiment and description thereof will consequently be omitted.

(3) Configuration and Function of Communicating Hole

A communicating hole 132 is formed in the sleeve 108, penetrating it axially. The region radially inward of the thrust bearing section 126, and the region formed by the lower-end face of the shaft 104, and the inner surfaces of the sleeve 108 and cover member 110 communicate by means of the communicating hole 132. Oil continuous with the oil retained within the series of bearing micro-gaps as described above is retained within the communicating hole 132. Here too the internal pressure of the oil retained within the communicating hole 132 balances with the internal pressure of the oil retained in the bearing sections. It will be appreciated that the functioning of the communicating hole 132 is likewise as is the case with its counterpart in the spindle motor of the foregoing first embodiment.

III. Disk-Drive Configuration

The internal configuration of a general disk-drive device 50 is represented in a schematic view in FIG. 8. A clean space where dust and debris are extremely slight is formed inside a housing 51, in the interior of which is installed a spindle motor 52 on which platter-shaped disks 53 for recording information are fitted. In addition, a head-shifting mechanism 57 that reads information from and writes information onto the disks 53 is disposed within the housing 51. The head-shifting mechanism 57 is constituted by: heads 56 that read/write information on the disks 53; arms 55 that support the heads 56; and an actuator 54 that shifts the heads 56 and arms 55 over the requisite locations on the disks 53.

Utilizing a spindle motor of the foregoing embodiments as the spindle motor 52 for the disk drive 50 as such enables the disk drive 50 to be made low-profile and reduced-cost, and at the same time improves the stability, reliability and endurance of the spindle motor to render a more highly reliable disk drive.

While single embodiments in accordance with the present invention of a spindle motor and a disk drive equipped therewith have been explained in the foregoing, the present invention is not limited to such embodiments. Various changes and modifications are possible without departing from the scope of the invention. 

1. A spindle motor comprising: a shaft; a unilaterally open-ended cylindrical bearing member having a bearing hole into which said shaft is inserted and a closed-end surface axially opposing the inserted-end face of said shaft; a rotor fixed to one end of said shaft for rotation together therewith, and having a circular flat face extending radially outward from the circumferential surface of said shaft; a serial bearing clearance filled with oil, formed in between the inner peripheral surface of said bearing hole in said bearing member, and the circumferential surface of said shaft, and in between the end face along the open end of said bearing member, and the flat face of said rotor; a thrust bearing section provided with dynamic-pressure-generating grooves contoured to impart on the oil pressure acting radially inward during rotation of said rotor, and formed in between the end face along the open end of said bearing member and the flat face of said rotor; a radial bearing section provided with dynamic-pressure-generating grooves contoured to impart on the oil pressure acting inward from both ends axially during rotation of said rotor, and formed in between the inner peripheral surface of said bearing hole and the circumferential surface of said shaft; and a communicating passage formed in said bearing member so that one end of said communicating passage opens on said thrust bearing section radially inward with respect to said dynamic-pressure-generating grooves therein and so that both axial ends of said bearing clearance where said clearance if formed in between the inner peripheral surface of said bearing hole and the circumferential surface of said shaft communicate through said passage, for balancing pressure within said bearing clearances.
 2. A spindle motor as set forth in claim 1, wherein: spiral grooves having a pump-in contour are provided as said dynamic-pressure-generating grooves in said thrust bearing section; said radial bearing section is axially separated twin constituents between the circumferential surface of said shaft and the inner peripheral surface of said bearing hole; and herringbone grooves in an axially unbalanced conformation are formed as said dynamic-pressure-generating grooves in at least either one of the twin constituents of said radial bearing section, for pressuring the oil toward the closed end of said bearing member from along the open end thereof.
 3. A spindle motor as set forth in claim 1, wherein: said bearing member is composed of a hollow cylindrical sleeve in which said bearing hole is provided, and a rimmed-cup-shaped, unilaterally closed-ended bearing housing for retaining said sleeve; said thrust bearing section is formed in between the end face along the rim of said bearing housing and the flat face of said rotor; and and the communicating passage is constituted by an axial groove formed in the circumferential surface of said sleeve, and by the inner peripheral surface of said housing.
 4. A spindle motor as set forth in claim 3, further comprising: a cylindrical wall provided on said rotor, depending from the flat face thereof and radially opposing across a gap the circumferential surface of said bearing housing; a tapered face that shrinks diametrically according as its outer diameter is away from the flat face of said rotor, provided on the circumferential surface of said bearing housing; wherein the oil forms and retains a meniscus between said tapered face and the inner peripheral surface of the cylindrical wall.
 5. A spindle motor as set forth in claim 4, wherein: a stepped portion is provided on said bearing housing where its circumferential surface continuous with said tapered face is recessed radially inward; a radially inward-projecting annular member corresponding to said stepped portion is affixed to the inner peripheral surface of said cylindrical wall on said rotor, and engagement between said stepped portion and said annular member constitutes a retainer for said rotor; and a micro-gap smaller than the minimum clearance dimension of the radial gap formed in between said tapered face of said bearing housing and the inner peripheral surface of said cylindrical wall on said rotor is formed in between an upper face, of said annular member and an undersurface of said stepped portion, functioning as a labyrinth seal.
 6. A spindle motor as set forth in claim 1, wherein: said bearing member is composed of a hollow cylindrical sleeve in which said bearing hole is formed, and one end of which is closed off by a cap member forming said closed-end surface; said thrust bearing section is formed in between the end face along the other end of said sleeve, and the flat face of said rotor; and said communicating passage is formed by a through-hole axially penetrating said sleeve.
 7. A spindle motor as set forth in claim 6, further comprising: a cylindrical wall provided on said rotor, depending from the flat face thereof and radially opposing across a gap the circumferential surface of said bearing housing; a tapered face that shrinks diametrically according as its outer diameter is away from the flat face of said rotor, provided on the circumferential surface of said sleeve; wherein the oil forms and retains a meniscus between said tapered face and the inner peripheral surface of the cylindrical wall.
 8. A spindle motor as set forth in claim 6, wherein: a stepped portion is provided on said sleeve where its circumferential surface continuous with said tapered face is recessed radially inward; a radially inward-projecting annular member corresponding to said stepped portion is affixed to the inner peripheral surface of said cylindrical wall on said rotor, and engagement between said stepped portion and said annular member constitutes a retainer for said rotor; and a micro-gap smaller than the minimum clearance dimension of the radial gap formed in between said tapered face of said sleeve and the inner peripheral surface of said cylindrical wall on said rotor is formed in between an upper face of said annular member and an undersurface of said stepped portion, functioning as a labyrinth seal.
 9. A spindle motor as set forth in claim 1, wherein said rotor is urged toward said closed-end surface of said bearing member by axially-acting magnetic attraction.
 10. A disk drive in which is mounted a disk-shaped recording medium onto which information is recordable, the disk drive including a housing; a spindle motor fixed within said housing, for spinning the recording medium; and an information accessing means for writing information into and reading information from requisite locations on said recording medium; the disk drive characterized in that said spindle motor comprises: a shaft; a unilaterally open-ended cylindrical bearing member having a bearing hole into which said shaft is inserted and a closed-end surface axially opposing the inserted end face of said shaft; a rotor fixed to one end of said shaft for rotation together therewith, and having a circular flat face extending radially outward from the circumferential surface of said shaft; a serial bearing clearance filled with oil, formed in between the inner peripheral surface of said bearing hole in said bearing member, and the circumferential surface of said shaft, and in between said bearing member along its open-end end face, and the flat face of said rotor; a thrust bearing section provided with dynamic-pressure-generating grooves contoured to impart on the oil pressure acting radially inward during rotation of said rotor, and formed in between the end face along the open end of said bearing member and the flat face of said rotor; a radial bearing section provided with dynamic-pressure-generating grooves contoured to impart on the oil pressure acting inward from both ends axially during rotation of said rotor, and formed in between the inner peripheral surface of said bearing hole and the circumferential surface of said shaft; and a communicating passage formed in said bearing member so that one end of said communicating passage opens on said thrust bearing section radially inward with respect to said dynamic-pressure-generating grooves therein and so that both axial ends of a one said bearing clearance where said clearance is formed in between the inner peripheral surface of said bearing hole and the circumferential surface of said shaft communicate through said passage, for balancing pressure within said bearing clearances.
 11. A disk drive as set forth in claim 10, wherein: spiral grooves having a pump-in contour are provided as said dynamic-pressure-generating grooves in said thrust bearing section; said radial bearing section is axially separated twin constituents between the circumferential surface of said shaft and the inner peripheral surface of said bearing hole; and herringbone grooves in an axially unbalanced conformation are formed as said dynamic-pressure-generating grooves in at least either one of the twin constituents of said radial bearing section, for pressuring the oil toward the closed end of said bearing member from along the open end thereof.
 12. A disk drive as set forth in claim 10, wherein: said bearing member is composed of a hollow cylindrical sleeve in which said bearing hole is provided, and a rimmed-cup-shaped, unilaterally closed-ended bearing housing for retaining said sleeve; said thrust bearing section is formed in between the end face along the rim of said bearing housing and the flat face of said rotor; and and the communicating passage is constituted by an axial groove formed in the circumferential surface of said sleeve, and by the inner peripheral surface of said housing.
 13. A disk drive as set forth in claim 12, further comprising: a cylindrical wall provided on said rotor, depending from the flat face thereof and radially opposing across a gap the circumferential surface of said bearing housing; a tapered face that shrinks diametrically according as its outer diameter is away from the flat face of said rotor, provided on the circumferential surface of said bearing housing; wherein the oil forms and retains a meniscus between said tapered face and the inner peripheral surface of the cylindrical wall.
 14. A disk drive as set forth in claim 13, wherein: a stepped portion is provided on said bearing housing where its circumferential surface continuous with said tapered face is recessed radially inward; a radially inward-projecting annular member corresponding to said stepped portion is affixed to the inner peripheral surface of said cylindrical wall on said rotor, and engagement between said stepped portion and said annular member constitutes a retainer for said rotor; and a micro-gap smaller than the minimum clearance dimension of the radial gap formed in between said tapered face of said bearing housing and the inner peripheral surface of said cylindrical wall on said rotor is formed in between an upper face of said annular member and an undersurface of said stepped portion, functioning as a labyrinth seal.
 15. A disk drive as set forth in claim 10, wherein: said bearing member is composed of a hollow cylindrical sleeve in which said bearing hole is formed, and one end of which is closed off by a cap member forming said closed-end surface; said thrust bearing section is formed in between the end face along the other end of said sleeve, and the flat face of said rotor; and said communicating passage is formed by a through-hole axially penetrating said sleeve.
 16. A disk drive as set forth in claim 15, further comprising: a cylindrical wall provided on said rotor, depending from the flat face thereof and radially opposing across a gap the circumferential surface of said bearing housing; a tapered face that shrinks diametrically according as its outer diameter is away from the flat face of said rotor, provided on the circumferential surface of said sleeve; wherein the oil forms and retains a meniscus between said tapered face and the inner peripheral surface of the cylindrical wall.
 17. A disk drive as set forth in claim 15, wherein: a stepped portion is provided on said sleeve where its circumferential surface continuous with said tapered face is recessed radially inward; a radially inward-projecting annular member corresponding to said stepped portion is affixed to the inner peripheral surface of said cylindrical wall on said rotor, and engagement between said stepped portion and said annular member constitutes a retainer for said rotor; and a micro-gap smaller than the minimum clearance dimension of the radial gap formed in between said tapered face of said sleeve and the inner peripheral surface of said cylindrical wall on said rotor is formed in between an upper face of said annular member and an undersurface of said stepped portion, functioning as a labyrinth seal.
 18. A disk drive as set forth in claim 10, wherein said rotor is urged toward said closed-end surface of said bearing member by axially-acting magnetic attraction. 